JPH11141468A - Displacement type fluid machinery - Google Patents

Abstract

PROBLEM TO BE SOLVED: To reduce the rotation moment acting on a rotating piston, dissolve the problem of friction and wear and improve the reliability by forming a part of the outer wall surface of a displacer which forms a contact when a space is formed by a cylinder inner wall and the displacer outer wall by the circumferential surface of a rolling element. SOLUTION: The inside form of a cylinder 4 in a displacement type compressing element housed in a sealed vessel and driven by a motor driven element is formed in such a manner that a hollow part has the same form every 120 deg.C. The end parts of the individual hollow parts form a plurality of substantially circular vanes 4b protruded inward (three in 3 row lap), and a rotating piston (displacer) 5 is arranged in such a hollow part with the mutual centers being shifted by ε so as to be engaged with the inside wall 4a and vane 4b of the cylinder 4. The cylinder 4 is formed so as to have an inside profile form in which an inner wall curve (g-a) is a spiral line having a winding angle of about 360 deg., and an outer wall curve (g-b) is a spiral line having a winding angle of about 360 deg..

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a pump, a compressor, an expander and the like, and more particularly to a positive displacement fluid machine.

[0002]

2. Description of the Related Art A reciprocating type fluid machine which moves a working fluid by repeating a reciprocating motion of a piston in a cylindrical cylinder has been used as a positive displacement type fluid machine for a long time, and a cylindrical piston is eccentric in a cylindrical cylinder. A rotary (rolling piston type) fluid machine that moves a working fluid by rotating, meshing a pair of fixed scroll and orbiting scroll having an upright spiral wrap on an end plate, and orbiting the orbiting scroll. 2. Description of the Related Art A scroll type fluid machine that moves a working fluid by a hydraulic fluid is known.

[0003] Reciprocating fluid machines have the advantage of being easy and inexpensive to manufacture because of their simple structure, but the stroke from the end of suction to the end of discharge is as short as 180 ° in terms of the rotation angle of the rotating shaft. The problem is that the flow velocity in the discharge process is high, the performance is reduced due to the increase in pressure loss, and the reciprocating motion of the piston makes it impossible to completely balance the unbalanced inertial force of the rotating shaft system. There is a problem that is large.

[0004] In the rotary type fluid machine, although the stroke from the end of suction to the end of discharge is 360 ° in the rotation angle of the rotating shaft, the problem that the pressure loss in the discharge process increases is smaller than that of the reciprocating type fluid machine. Since the gas is discharged once per rotation of the shaft, the fluctuation of the gas compression torque is relatively large, and there is a problem of vibration and noise as in the reciprocating fluid machine.

Further, in the scroll type fluid machine, the stroke from the end of the suction to the end of the discharge requires a rotation angle of the rotary shaft of 360 °.
The above is a long time (usually 9
(Approximately 00 °), there is an advantage that the pressure loss in the discharge process is small, and since a plurality of working chambers are generally formed, the fluctuation of the gas compression torque during one rotation is small and vibration and noise are small.

However, it is necessary to manage the clearance between the spiral wraps in the lap meshing state and the clearance between the end plate and the tip of the lap, so that high-precision processing must be performed and the processing cost is high. Problem. In addition, there is a problem that the stroke from the end of suction to the end of discharge is as long as 360 ° or more in the rotation angle of the rotating shaft, and the longer the period of the compression process, the more internal leakage increases.

By the way, the displacer (swirl piston) for moving the working fluid does not rotate relative to the cylinder into which the working fluid is sucked, but revolves around a substantially constant radius, that is, swirls. Japanese Patent Application Laid-Open No. 55-23353 (Reference 1), US Pat. No. 2,112,890 (Reference 2), and Japanese Patent Application Laid-Open No. 5-2
It is proposed in Japanese Patent Publication No. 02869 (Reference 3) and Japanese Patent Application Laid-Open No. 6-280758 (Reference 4). The displacement type fluid machine proposed here comprises a piston having a petal shape in which a plurality of members (vanes) extend radially from the center, and a cylinder having a hollow portion substantially similar to the piston, This piston moves the working fluid by revolving in the cylinder.

[0008]

Problems to be Solved by the Invention Documents 1 to 4 mentioned above
Since the positive displacement type fluid machine shown in (1) does not have a reciprocating portion unlike the reciprocating type, it is possible to balance the imbalance of the rotating shaft system. Therefore, vibration is small,
Furthermore, since the relative sliding speed between the piston and the cylinder is low, the friction loss can be relatively reduced.

However, the rotation angle of the rotating shaft from the end of suction to the end of discharge of each working chamber is small, and the time required from the end of discharge of the working fluid to the start of the next (compression) stroke (end of suction). There is a shift (time lag), and the working chamber from the end of suction to the end of discharge is formed to be biased around the rotation axis. As a reaction force, a rotational moment for rotating the piston itself acts excessively on the piston, and there is a disadvantage that reliability problems such as friction and wear of the vane easily occur.

SUMMARY OF THE INVENTION It is an object of the present invention to provide a positive displacement type fluid machine which can reduce the rotation moment acting on the orbiting piston and solve the problems of friction and wear.

[0011]

Means for achieving the above object include the following modes.

(1) A cylinder having an inner wall formed between a pair of end plates and having a continuous curved surface, and an outer wall provided so as to face the inner wall of the cylinder. The outer wall and the end plate are provided with a displacer forming a plurality of spaces, and the spaces present along the contour of the displacer are alternately a space in a suction stroke and a space in a compression or discharge stroke. In the displacement type fluid machine in which the cylinder inner wall and the displacer outer wall curve are formed, a part of the outer wall surface of the displacer serving as a contact point when a space is formed by the cylinder inner wall and the displacer outer wall is formed on the outer peripheral surface of the rolling element. Formed.

(2) A cylinder having an inner wall formed between the end plates and having a continuous curved plane, and an outer wall provided so as to face the inner wall of the cylinder. The outer wall and the end plate are provided with a displacer forming a plurality of spaces, and the spaces present along the contour of the displacer are alternately a space in a suction stroke and a space in a compression or discharge stroke. In the displacement type fluid machine in which the cylinder inner wall and the displacer outer wall curve are formed, a part of the inner wall surface of the cylinder, which is a contact point when a space is formed by the cylinder inner wall and the displacer outer wall, is formed on the outer peripheral surface of the rolling element. Formed.

(3) A cylinder having an inner wall formed between the end plates and having a continuous curved shape in a plane, and an outer wall provided so as to face the inner wall of the cylinder. The outer wall and the end plate are provided with a displacer forming a plurality of spaces, and the spaces present along the contour of the displacer are alternately a space in a suction stroke and a space in a compression or discharge stroke. In the displacement type fluid machine in which the cylinder inner wall and the displacer outer wall curve are formed, a part of the outer wall surface of the displacer and a part of the cylinder inner wall which serve as a contact point when a space is formed by the cylinder inner wall and the displacer outer wall. The part is formed on the outer peripheral surface of the rolling element.

(4) A cylinder having an inner wall formed between the end plates and having a continuous curved plane, and an outer wall provided so as to face the inner wall of the cylinder. The outer wall and the end plate are provided with a displacer forming a plurality of spaces, and the spaces present along the contour of the displacer are alternately a space in a suction stroke and a space in a compression or discharge stroke. In the displacement type fluid machine in which the cylinder inner wall and the displacer outer wall curve are formed, a part of the outer wall surface of the displacer or a part of the inner wall surface of the cylinder which is a contact point when a space is formed by the inner wall of the cylinder and the outer wall of the displacer is formed. Formed on the outer peripheral surface of the rolling element,
And performing a surface treatment to form a conformable film on the outer peripheral surface of the displacer or the inner wall surface of the cylinder to which the rolling elements are attached, so that the outer peripheral surface of the rolling elements and the outer peripheral surface of the displacer or the inner surface of the cylinder are formed. The minimum distance t from the wall surface is 0>t> -T _st (T _st : surface treatment film thickness, minus (-) when the outer peripheral surface of the displacer or the inner wall surface of the cylinder protrudes from the outer peripheral surface of the rolling element. A) Install the rolling element at the position where

[0016]

DESCRIPTION OF THE PREFERRED EMBODIMENTS The features of the present invention described above will be further clarified by the following embodiments. Hereinafter, an embodiment of the present invention will be described with reference to the drawings. First, the structure of a swirling type fluid machine according to an embodiment of the present invention will be described with reference to FIGS. FIG. 1A is a vertical cross-sectional view (A-A cross-sectional view of FIG. 1B) showing a main part of a hermetic compressor when a positive displacement fluid machine according to an embodiment of the present invention is used as a compressor. FIG. 2B is a plan view showing a state in which a compression chamber is formed as viewed from arrows BB in FIG. 2A, FIG. 2 is an operation principle diagram of a positive displacement compression element, and FIG. 3 is an embodiment of the present invention. It is a longitudinal section of a hermetic compressor when a positive displacement type fluid machine is used as a compressor.

In FIG. 1, a positive displacement element 1 and a motor-driven element 2 for driving the same are provided in a closed container 3 (not shown).
Is stored. The details of the displacement type compression element 1 will be described. FIG. 1B shows a three-row wrap in which three sets of the same contour shape are combined. The inner peripheral shape of the cylinder 4 is
The hollow portion is formed such that the same shape appears every 120 ° (center o ′). At the end of each individual hollow,
It has a plurality of (in this case, three wraps, there are three) quasi-arc shaped vanes 4b protruding inward. The orbiting pistons 5 are arranged inside the cylinder 4 and are offset from each other by ε so that they are engaged with the inner peripheral wall 4a (the portion having a larger curvature than the vane 4b) of the cylinder 4 and the vane 4b. When the center o 'of the cylinder 4 matches the center o of the revolving piston 5, a gap having a constant width is formed as a basic shape between the two contours.

Next, the operation principle of the displacement type compression element 1 is shown in FIG.
And FIG. Symbol o is the center of the revolving piston 5 which is a displacer, and symbol o 'is the center of the cylinder 4 (or the drive shaft 6). The symbols a, b, c, d,
Symbols e and f denote a point of contact between the inner peripheral wall 4a and the vane 4b of the cylinder 4 and the revolving piston 5. Here, looking at the inner peripheral contour shape of the cylinder 4, the combination of the same curves is 3
The parts are connected smoothly and continuously. Focusing on one of these, the curve forming the inner peripheral wall 4a and the vane 4b can be regarded as one thick vortex curve (the tip of the vane 4b is considered to be the start of the vortex), and the inner wall curve (G
-A) means that the winding angle, which is the sum of the respective arc angles forming the curve, is approximately 360 ° (the design concept is 360 °, but this value is not exactly the same due to manufacturing errors. The same applies hereinafter.). The details of this winding angle will be described later.)
The outer wall curve (gb) has a winding angle of about 360
° vortex curve.

As described above, the one inner peripheral contour shape is formed from the inner wall curve and the outer wall curve. The two spiral curves are arranged at substantially equal pitches (120 ° because of the three-line wrap), and the outer wall curve and the inner wall curve of the adjacent spiral body are smoothly connected to each other by a smooth connection curve (bb ′) such as an arc. ), The entire inner peripheral contour shape of the cylinder 4 is formed. The outer peripheral shape of the revolving piston 5 is also configured according to the same principle as that of the cylinder 4.

The spiral body composed of three curves is arranged at substantially equal pitches (120 °) on the circumference. This is for the purpose of uniformly distributing the load accompanying the compressing operation described later and for the purpose of manufacturing. In consideration of easiness, especially when these are not a problem, the pitch may be unequal.

Now, the compression operation by the cylinder 4 and the revolving piston 5 configured as described above will be described with reference to FIG. Reference numeral 7a denotes a suction port, and 8a denotes a discharge port, which are provided on end plates respectively corresponding to three places. By rotating the drive shaft 6, the orbiting piston 5 revolves around the orbital radius ε (= oo ') without rotating around the center o' of the cylinder 4 on the fixed side, and the orbiting piston 5
Of the plurality of working chambers 15 (in a plurality of hermetically sealed spaces surrounded by the inner peripheral contour (inner wall) of the cylinder and the outer peripheral contour (side wall) of the piston), the suction (compression) process is completed. In other words, the space is the period from the end of suction to the end of discharge.If the above-mentioned winding angle is limited to 360 °, this space disappears at the end of compression, but the suction ends at that moment. Therefore, this space is counted as one.However, when used as a pump,
A space communicating with the outside via the discharge port is formed (in this embodiment, three working chambers are always present). One working chamber surrounded by contact points a and b and hatched (separated into two at the end of suction, but immediately after the compression stroke is started, these two working chambers are connected and become one) ) Will be described.

FIG. 2A shows a state in which the suction of the working gas from the suction port 7a into this working chamber has been completed. FIG. 2B shows a state in which the drive shaft 6 rotates 90 ° from this state, and FIG. 2C shows a state where the drive shaft 6 rotates 180 ° from the beginning.
FIG. 2D shows a state in which the rotation is further advanced and the rotation is performed by 270 ° from the beginning. When rotated 90 ° from FIG. 2 (4), it returns to the initial state of FIG. 2 (1). Thus, as the rotation proceeds, the working chamber 15 reduces its volume and discharge port 8a.
Is closed by the discharge valve 9 (shown in FIG. 1), so that the working fluid is compressed.

When the pressure in the working chamber 15 becomes higher than the external discharge pressure, the discharge valve 9 automatically opens due to the pressure difference, and the compressed working gas is discharged through the discharge port 8a. The rotation angle of the rotating shaft from the end of the suction (compression start) to the end of the discharge is 360 °, and the next suction stroke is prepared during each of the compression and discharge strokes. Starts compression. For example, focusing on the space formed by the contacts a and d, the suction has already been started from the suction port 7a at the stage of FIG. 2A, and the volume increases as the rotation progresses. When the state is reached, this space is divided. Fluid corresponding to this divided amount is supplemented from the space formed by the contacts b and e.

The manner of supplementation will be described in detail. FIG.
In the space formed by the contacts a and d adjacent to the working chamber formed by the contacts a and b in the state (1), suction has started. This space is once expanded as shown in FIG. 2 (3), and then divided when it becomes FIG. 2 (4). Therefore,
Not all the fluid in the space formed by the contacts a and d is compressed in the space formed by the contacts a and b. 2 (4), the space formed by the contacts b and e in the suction process in FIG. 2 (4) is the same as the fluid volume that has been divided and not taken into the space formed by the contacts a and d. As shown in 1), it is divided and filled with the fluid flowing into the space formed by the contacts e and b near the discharge port.

This is because, as described above, the wraps are arranged at a uniform pitch. That is, since the shapes of the revolving piston and the cylinder are formed by repetition of the same contour shape, substantially the same amount of fluid can be compressed even if any working chamber obtains fluid from different spaces. Although it is possible to perform processing so that the volumes formed in the respective spaces are equal even if the pitch is not uniform, the productivity is poor. In any of the above prior arts, the space in the suction process is closed and the internal fluid is compressed and discharged as it is, whereas the space in the suction process adjacent to the working chamber is divided in this way. Performing the compression operation is one of the features of the present embodiment.

As described above, the working chambers for the continuous compression operation are distributed at substantially equal pitches around the drive bearing 5a located at the center of the revolving piston 5, and each working chamber is The compression is performed out of phase. That is, 1
Focusing on the two spaces, the rotation angle of the rotating shaft is 360 ° from the suction to the discharge, but in the present embodiment, three working chambers are formed and these discharge at a phase shifted by 120 °. When the compressor is operated as a compressor that compresses the refrigerant, the refrigerant is discharged three times during a rotation angle of the rotating shaft of 360 °.

Assuming that the space at the moment when the compression operation is completed (the space surrounded by the contact points a and b) is one space, if the winding angle is 360 ° as in this embodiment, any compression Even in the machine operating state, the space in the suction stroke and the space in the compression stroke are designed to be alternated, so that the next compression stroke is started immediately after the compression stroke is completed. Fluid can be compressed smoothly and continuously.

Next, the volume having such a shape (turning)
A compressor incorporating the mold compression element 1 will be described with reference to FIGS. In FIG. 3, in addition to the above-described cylinder 4 and the revolving piston 5, a revolving-type compression element 1 includes a drive shaft that drives the revolving piston 5 by fitting a crank part 6 a to a bearing at the center of the revolving piston 5. 6, a main bearing 7 and an auxiliary bearing 8, which also serve as an end plate for closing both end openings of the cylinder 4 and a bearing for supporting the drive shaft 6, a suction port 7a formed in an end plate of the main bearing 7, It has a discharge port 8a formed on the end plate of the sub bearing 8, and a discharge valve 9 for opening and closing the discharge port 8a with a differential pressure. However, the discharge valve 9 may be of a reed valve type. 5 b is a through hole formed in the revolving piston 5. Reference numeral 10 denotes a suction cover attached to the main bearing 7, and reference numeral 11 denotes a discharge cover for integrally forming the discharge chamber 8b with the sub bearing 8.

The electric element 2 includes a stator 2a and a rotor 2b, and the rotor 2b is fixed to the drive shaft 6 by shrink fitting or the like. The electric element 2 is constituted by a brushless motor for improving electric motor efficiency, and is driven and controlled by a three-phase inverter. However, 2 may be another motor type, for example, a DC motor or an induction motor.

Numeral 12 denotes lubricating oil stored at the bottom of the sealed container 3, in which the lower end of the drive shaft 6 is immersed.
Reference numeral 13 denotes a suction pipe, reference numeral 14 denotes a discharge pipe, and reference numeral 15 denotes the above-described working chamber formed by the engagement between the inner peripheral wall 4a and the vane 4b of the cylinder 4 and the revolving piston 5. Also,
The discharge chamber 8b is separated from the pressure in the closed container 3 by a seal member 16 such as an O-ring.

The flow of working gas (refrigerant gas) when the positive displacement fluid machine in this embodiment is used as an air conditioning compressor will be described with reference to FIG. As indicated by the arrow in the drawing, the working gas that has entered the closed casing 3 through the suction pipe 13 enters the suction cover 10 attached to the main bearing 7, and passes through the suction port 7a to move the positive displacement element 1 Then, the rotation of the drive shaft 6 causes the revolving piston 5 to perform a revolving motion to reduce the volume of the working chamber, thereby being compressed. The compressed working gas pushes up the discharge valve 9 through the discharge port 8 a formed in the end plate of the sub-bearing 8, enters the discharge chamber 8 b, and flows out through the discharge pipe 14 to the outside. The reason why the gap is formed between the suction pipe 13 and the suction cover 10 is to cool the electric motor element by flowing the working gas into the electric motor element 2 as well.

The lubricating oil stored inside is sent to each sliding portion from the bottom through a hole provided inside the drive shaft and lubricated by differential pressure or centrifugal pump oil supply. This part is also supplied to the inside of the working chamber through the gap.

Here, an example of a method of forming the contour shapes of the revolving piston 5 and the cylinder 4 which are the main parts constituting the positive displacement compression element 1 of the present invention will be described with reference to FIGS. Is taken as an example). FIG. 4 (a) (b)
Is an example of the shape of a revolving piston whose planar shape is formed by a combination of arcs as an example, (a) is a plan view,
(B) is a side view. 5 (a) and 5 (b) are examples of a cylinder shape which meshes with the pair of the revolving pistons shown in FIG. 4, where (a) is a plan view and (b) is a side view. Also,
FIG. 6 is a diagram in which the center o of the revolving piston shown in FIG. 4 and the center o ′ of the cylinder shown in FIG.

In FIG. 4 (a), in the plane shape of the revolving piston, three identical contours are continuously connected around the center o (center of the equilateral triangle IJK). The contour shape is formed by a total of seven arcs from a radius R1 to a radius R7, and points p, q, r, s, t, u, v, and w are connection points of arcs having different radii. . The curve pq is an arc of radius R1 centered on one side IJ of the equilateral triangle, where point p is at a distance R7 from vertex I. The curve qr is an arc of radius R2 having a center on an extension of a straight line connecting the contact point q and the center of the radius R1, and the curve rs is an arc of radius R3 having a center on a straight line connecting the contact point r and the center of the radius R2. Is an arc of a radius R4 having a center on an extension of a straight line connecting the contact s and the center of the radius R3.

The curve tu is an arc of a radius R5 having a center on an extension of a straight line connecting the contact point t and the center of the radius R4, and a curve uv.
Is an arc of radius R6 centered on a centroid o on an extension of a straight line connecting the contact point u and the center of the radius R5, and a curve vw is a vertex J on a straight line connecting the contact point v and the center of the radius R6 (centroid o). This is an arc having a radius R7 at the center. Note that the radii R1, R2, R
The angles of the respective arcs of R3, R4, R5, and R6 are determined by the condition that the contacts are smoothly connected (the inclination of the tangent line at the contacts is the same).

When the contour from point p to point w is rotated 120 ° counterclockwise around the center of gravity o, point p overlaps point w, and when further rotated by 120 °, the contour around the entire circumference is completed. . As a result, the plane shape of the revolving piston is obtained,
By providing the thickness h, a revolving piston is formed.

When the plane shape of the revolving piston is determined, when the revolving piston makes a revolving motion with a revolving radius ε, the contour of the cylinder that meshes with the revolving piston becomes a curve forming the contour of the revolving piston as shown in FIG. Is an offset curve of ε.

The outline shape of the cylinder will be described with reference to FIG. The triangle IJK is the same equilateral triangle as in FIG. The contour shape is formed by a total of seven circular arcs like the revolving piston, and the points p ′, q ′, r ′, s ′, t ′, u ′,
v 'and w' are connection points of arcs having different radii. A curve p′q ′ is an arc having a radius (R1−ε) centered on one side IK of an equilateral triangle.
7 + ε). The curve q'r 'is an arc of a radius (R2-?) Having a center on an extension of a straight line connecting the contact q' and the center of the radius (R1-?), And the curve r's 'is a contact r' and a radius (R2 -[Epsilon]) an arc having a radius (R3- [epsilon]) centered on a straight line connecting the center, and a curve s't 'is a radius having a center on a straight line connecting s' and the center of the radius (R3- [epsilon]) (R
4 + ε).

The curve t'u 'has a contact point t' and a radius (R4 +
radius (R) having a center on an extension of a straight line connecting the centers of ε)
5 + ε), the curve u′v ′ has a contact point u ′ and a radius (R5
+ Ε) is an arc of a radius (R6 + ε) centered on the centroid o ′ on an extension of a straight line connecting the centers of the points V ′ and w ′
And an arc having a radius (R7 + ε) centered on a vertex J on a straight line connecting the center (center o ′) of the radius (R6 + ε).
Note that the radii (R1-ε), (R2-ε), (R3-ε),
The angle of each of the arcs of (R4 + ε), (R5 + ε), and (R6 + ε) is determined by the condition that the contact points are smoothly connected at the respective contact points (the inclination of the tangent line at the contact points is the same) as in the case of the revolving piston.

When the contour shape from the point p 'to the point w' is rotated counterclockwise about the center o 'by 120 °, the point p' coincides with the point w '. The contour shape is completed. Thereby, the planar shape of the cylinder is obtained. The thickness H of the cylinder is slightly larger than the thickness h of the revolving piston.

FIG. 6 is a diagram showing a part of the center of the revolving piston overlapped with the center of the cylinder o '. The gap formed between the turning piston and the cylinder is set to ε equal to the turning radius. It is desirable that this gap be ε over the entire circumference. However, as long as the working chamber formed by the outer peripheral contour of the revolving piston and the inner peripheral contour of the cylinder operates normally, for some reason,
There may be places where this relationship breaks.

Here, the method of combining the plurality of arcs has been described as a method of forming the contours of the outer wall of the revolving piston and the inner wall of the cylinder. However, the present invention is not limited to this, and the present invention is not limited to this. A similar contour shape can be formed by a combination of curves.

The operation and effect of the embodiment described with reference to FIGS. 1 to 6 will be described below. FIG. 7 shows the volume change characteristic of the working chamber according to the present invention (represented by the ratio between the suction volume Vs and the working chamber volume V) with the rotation angle θ of the rotating shaft from the end of suction as the horizontal axis.
Are shown in comparison with other types of compressors. Accordingly, the volume change characteristic of the positive displacement type compression element 1 according to the present embodiment is based on a kind of operating condition of the air conditioner having the discharge start volume ratio of 0.37 (for example, when the working gas is Freon HCFC22, the suction pressure Ps = 0.64).
Comparing the pressure and discharge pressure Pd = 2.07 MPa), the compression process is almost the same as that of the reciprocating type, and the compression process is completed in a short time, so that leakage of working gas is reduced and the capacity and efficiency of the compressor are reduced. Can be improved. On the other hand, the discharge process is about 50 times faster than the rotary type (rolling piston type).
%, The pressure loss is reduced because the discharge flow rate is reduced, and the fluid loss (excessive compression loss) in the discharge process can be greatly reduced to improve the performance.

FIG. 8 shows a change in the amount of work during one rotation of the rotating shaft, that is, a change in the gas compression torque T in this embodiment, in comparison with other types of compressors (where Tm is an average torque). ). Thus, the torque fluctuation of the displacement type compression element 1 of the present invention is very small, about 1/10 of that of the rotary type, and is equivalent to that of the scroll type. Since there is no mechanism for performing the rotation, the vibration and noise of the compressor in which the inertia of the rotary shaft system is balanced can be reduced.

Further, as shown in FIG. 4, since the contour is not a long spiral shape like a scroll type, the processing time and cost can be reduced, and there is no end plate (end plate) for maintaining the spiral shape. Therefore, it can be manufactured by processing similar to a rotary type as compared with a scroll type in which processing cannot be performed by penetrating a jig.

Further, since the thrust load due to the gas pressure does not act on the orbiting piston, it is easy to manage the axial clearance which significantly affects the performance of the compressor as seen in the scroll compressor, so that the performance can be improved. Furthermore, as a result of the calculation, the thickness can be reduced as compared with a scroll compressor having the same volume and the same outer diameter,
It can also contribute to reducing the size and weight of the compressor.

Next, the relationship between the above-mentioned winding angle and the rotation angle θc of the rotating shaft from the end of suction to the end of discharge will be described.
In the above-described embodiment, the winding angle is described as 360 °. However, it is also possible to change the rotation angle θc of the rotating shaft by changing the winding angle. For example, in FIG. 2, since the winding angle is 360 °, the rotation angle θc of the rotating shaft from the end of suction to the end of discharge returns to the original state when the rotation angle θc is 360 °. When the rotation angle θc of the rotating shaft from the end of suction to the end of discharge is reduced by making the winding angle smaller than 360 °, a state occurs in which the discharge port communicates with the suction port, and the expansion action of the fluid in the discharge port occurs. This causes a problem that the fluid once sucked flows backward.

When the winding angle is larger than 360 °, the rotation angle of the rotating shaft is also larger than 360 °, and two working chambers having different sizes are formed from the end of suction to the communication with the space having the discharge port. You. When this is used as a compressor, irreversible mixing loss occurs when the two working chambers merge because the pressure rises of these two working chambers are different from each other, resulting in an increase in compression power. Further, even if it is used as a liquid pump, it is difficult to apply it as a pump because an operation chamber that does not communicate with the discharge port is formed. For this reason, it can be said that the winding angle is desirably 360 ° as much as possible within the range of allowable accuracy.

The above-mentioned JP-A-55-23353 (Reference 1)
Rotation angle θ of the rotary shaft during the compression stroke in the fluid machine described in
c is 180 °, and the rotation angle θc of the rotary shaft during the compression stroke in the fluid machine described in JP-A-5-202869 (Reference 3) and JP-A-6-280758 (Reference 4).
Is θc = 210 °. During the period from the end of discharge of the working fluid to the start of the next compression stroke (end of suction),
In Reference 1, the rotation angle θc of the rotating shaft is 180 °, and in References 3 and 4, it is 150 °.

When the rotation angle θc of the rotating shaft in the compression stroke is 210 °, each working chamber (reference numerals I, II,
FIG. 9 (a) shows a compression stroke diagram of the compression stroke (indicated by III and IV).
However, the number of rows N = 4. The rotation angle θc of the rotation shaft is 360
Although four working chambers are formed in ゜, the number n of working chambers simultaneously formed at a certain angle is n = 2 or 3. The maximum value of the number of working chambers formed simultaneously is 3, which is smaller than the number of working chambers.

Similarly, FIG. 10A shows the case where the number of threads N = 3 and the rotation angle θc of the rotating shaft in the compression stroke is 210 °. Also in this case, the number n of working chambers formed simultaneously is n = 1.
Alternatively, it is 2, and the maximum value of the number of working chambers formed simultaneously is 2, which is smaller than the number of rows.

In such a state, the working chamber is formed so as to be deviated around the drive shaft, so that a dynamic imbalance occurs, and the rotation moment acting on the orbiting piston becomes excessive.
The contact load between the orbiting piston and the cylinder increases, and there is a problem that the performance decreases due to an increase in mechanical friction loss and the reliability decreases due to vane wear.

In order to solve this problem, in the present embodiment, the rotation angle θc of the rotating shaft from the end of suction to the end of discharge (sometimes called a compression stroke) is

[0054]

[Formula 1] (((N-1) / N) · 360 °) <θc ≦ 360 ° (Formula 1) An outer contour shape of the revolving piston and an inner contour shape of the cylinder are formed so as to satisfy I have.

In other words, the above-mentioned winding angle is in the range of Expression 1. Referring to FIG. 9B, the rotation angle θc of the rotating shaft in the compression stroke is larger than 270 °, and the number n of working chambers simultaneously formed is n = 3 or 4, and the maximum number of working chambers is The value is 4. This value is the number of articles N
(= 4). In FIG. 10B, the rotation angle θc of the rotation shaft in the compression stroke is larger than 240 °, and the number n of working chambers formed simultaneously is n = 2 or 3
And the maximum value of the number of working chambers is 3. This value is equal to the number N of rows (= 3).

As described above, the rotation angle θc of the rotating shaft in the compression stroke
Is larger than the value on the left side of Equation 1, the maximum value of the number of working chambers becomes equal to or more than the number N, and the working chambers are dispersedly arranged around the drive shaft. As a result, the rotational moment acting on the orbiting piston is reduced, the contact load between the orbiting piston and the cylinder is reduced, and the reliability of the contact portion can be improved as well as the performance is improved by reducing the mechanical friction loss.

On the other hand, the upper limit of the rotation angle θc of the rotating shaft in the compression stroke is 360 ° according to the equation (1). The upper limit of the rotation angle θc of the rotation shaft in this compression stroke is 360 °. As described above, the time lag from the end of the discharge of the working fluid to the start of the next compression stroke (end of suction) can be made zero, and the gas in the gap volume occurring when θc <360 ° can be re-established. It is possible to prevent a reduction in suction efficiency due to expansion, and to prevent irreversible mixing loss that occurs when the two working chambers merge due to different pressure increases in the two working chambers when θc> 360 °. The latter will be described with reference to FIG.

FIG. 11 shows a positive displacement type fluid machine in which the compression stroke is 375 ° at the rotation angle θc of the rotating shaft. FIG. 11 (a)
5 shows a state in which the suction of the two working chambers 15a and 15b has been completed. At this time, the pressures of the two working chambers 15a and 15b are equal at the suction pressure Ps. Discharge port 8a
Is located between the working chambers 15a and 15b and is not in communication with both working chambers. From this state, the rotation angle θc of the rotation shaft
FIG. 11B shows a state in which the rotation has been advanced by 15 °. This is a state immediately before the discharge port 8a and both working chambers 15a and 15b communicate with each other. At this time, the volume of the working chamber 15a is smaller than that at the end of the suction in FIG. 11A, and the compression is progressing, and the pressure is higher than the suction pressure Ps. On the contrary,
Conversely, the volume of the working chamber 15b is larger than at the end of the suction, and the pressure is also lower than the suction pressure Ps due to the expansion action.

When the next instantaneous working chambers 15a and 15b are united (communicated), irreversible mixing as shown by an arrow in FIG. 11C occurs, and the performance decreases due to an increase in compression power. Become. Therefore, the upper limit of the rotation angle θc of the rotation shaft in the compression stroke is desirably 360 °.

FIG. 12 shows a compression element of a positive displacement fluid machine described in Document 3 or Document 4, wherein (a) is a plan view,
(B) is a side view. The number of threads N is 3, and the rotation angle θc (winding angle θ) of the rotation shaft in the compression stroke is 210 °. In this figure, the number n of working chambers is n = 1 or 2, as shown in FIG. This figure shows the rotation angle θ of the rotating shaft.
Indicates a state of 0 °, and the number n of working chambers is 2. As is clear from this figure, the space on the right side of the space formed by the outer peripheral contour of the revolving piston and the inner peripheral contour of the cylinder does not serve as a working chamber, but instead has a suction port 7 a and a discharge port 8.
a communicates. Therefore, the gas once flowing into the cylinder 4 from the suction port 8a flows backward due to the re-expansion of the gas in the gap volume of the discharge port 7a, and there is a problem that the suction efficiency is reduced.

Now, consider a case where the rotation angle θc of the rotation shaft in the compression stroke of the positive displacement type fluid machine shown in FIG. 12 is expanded using the concept of the present embodiment. In order to increase the rotation angle θc of the rotation shaft during the compression stroke, the winding angle of the contour curve of the cylinder 4 must be increased as shown by a two-dot chain line. It is difficult to make the rotation angle θc of the rotation shaft of the compression stroke larger than 240 ° so that the thickness becomes thin and the maximum value of the number n of the working chambers becomes more than the number N of rows (N = 3).

FIG. 13 shows an example of a compression element of a displacement type fluid machine having the same stroke volume (suction volume), the same outer diameter, and the same turning radius as the displacement type fluid machine shown in FIG. The rotation angle θc of the rotation shaft in the compression stroke of the compression element shown in FIG. 13 realizes 360 ° larger than 240 °. This is because, in the compression element shown in FIG. 12, the gap between the seal points forming the working chamber is formed by a smooth curve, and therefore, for example, the rotation of the rotation shaft during the compression stroke is performed based on the concept of the present embodiment. Although a maximum of 240 ° is the limit for increasing the angle θc, in the compression element according to the present embodiment shown in FIG. 13, the distance between the sealing points (ac) is not smooth (in the case of a uniform curve). However, the shape near the contact point b is formed so as to protrude when viewed from the revolving piston, and there is a constriction on the way from the center of the revolving piston toward the tip.

These are also applicable to the embodiment shown in FIG. With these shapes, the winding angle from the contact a to the contact b can be set to 360 ° larger than 240 °, and the winding angle from the contact b to the contact c can be set to 240 °.
It can be greater than {360}. As a result,
The rotation angle θc of the rotation shaft in the compression stroke should be
60 °, and the maximum value of the number n of working chambers is
The above can be considered. For this reason, the working chambers are dispersed and the rotation moment can be reduced.

Further, when the cylinder height (thickness) of the compression element shown in FIG. 12 is set to H due to the increase in the number of working chambers which can function effectively, the compression element shown in FIG. Since the cylinder height is 0.7H, which is 30% lower, the size of the compression element can be reduced.

FIG. 14 is an explanatory diagram of the load and moment acting on the revolving piston 5 in the present embodiment. The symbol θ is the rotation angle of the drive shaft 6, and ε is the turning radius. With the compression of the working gas, a tangential force Ft perpendicular to the eccentric direction and a radial force Fr corresponding to the eccentric direction act on the orbiting piston 5 due to the internal pressure of each working chamber 15 as shown in the figure.
The resultant force of Ft and Fr is F. A rotation moment M (= F · F) for rotating the revolving piston due to the deviation of the resultant force F from the center o of the revolving piston 5 (length of the arm 1).
l) works. This rotation moment M is supported by the reaction force R at the contact point g and the contact point b of the orbiting piston 5 and the cylinder 4.
1 and the reaction force R2.

According to the present invention, at all times, 2 near the suction port 7a
In addition, a moment is received at three or three contact points, and no reaction force acts on the other contact points. The displacement type compression element 1 of the present invention
Around the crank 6a of the drive shaft 6 fitted to the center of the revolving piston 5, working chambers in which the rotation angle of the rotating shaft from the end of suction to the end of discharge becomes approximately 360 ° at a substantially equal pitch are distributed. With this arrangement, the point of action of the resultant force F can be made closer to the center o of the revolving piston 5, and the length l of the arm of the moment can be reduced to reduce the rotation moment M. Therefore, the reaction force R1 and the reaction force R2 are reduced.
Further, as can be seen from the positions of the contact point g and the contact point b, the sliding portion between the revolving piston 5 and the cylinder 4 receiving the rotation moment M is moved to the working gas inlet 7a having a low temperature and a high oil viscosity.
Since it is set to be in the vicinity, it is possible to provide a positive displacement fluid machine in which an oil film of the sliding portion is easily secured.

FIG. 15 shows a comparison between the compression element shown in FIG. 12 and the compression element shown in FIG. 13 in terms of the rotational moment M during one rotation of the shaft acting on the orbiting piston due to the internal pressure of the working fluid. The calculation conditions are the refrigeration conditions of the working fluid HFC134a (suction pressure Ps = 0.095 MPa, discharge pressure Ps
d = 1.043 MPa). Thus, the number of working chambers n
In the compression element according to the present embodiment in which the maximum value is equal to or greater than the number of lines, the working chambers from the end of suction to the end of discharge are dispersed and arranged at substantially equal pitches around the drive shaft, so that the mechanical balance is good. That is, the load vector due to the compression can be configured to be substantially directed to the center. Therefore, the rotation moment M acting on the orbiting piston can be reduced. As a result,
The contact load between the revolving piston and the cylinder is reduced, so that the mechanical efficiency can be improved and the reliability as a compressor can be improved.

Here, the relationship between the period during which the suction port 7a and the discharge port 8a communicate with each other and the rotation angle of the compression stroke rotating shaft will be described. A time lag Δθ represented by the rotation angle of the rotating shaft between the time when the suction port and the discharge port communicate with each other, that is, from the end of discharge of the working fluid to the start of the next compression stroke (end of suction).
Is Δθ = 36 as the rotation angle θc of the rotation shaft in the compression stroke.
0 ° −θc.

When Δθ ≦ 0 °, there is no period in which the suction port and the discharge port communicate with each other, so that the suction efficiency does not decrease due to the re-expansion of the gas in the gap volume of the discharge port.

When Δθ> 0 °, there is a period in which the suction port and the discharge port communicate with each other, so that the suction efficiency is reduced due to the re-expansion of the gas in the gap volume of the discharge port, and
(Refrigeration) capacity will be reduced. Further, a decrease in the suction efficiency (volume efficiency) leads to a decrease in the adiabatic efficiency, which is the energy efficiency of the compressor, or the coefficient of performance.

The rotation angle θc of the rotating shaft in the compression stroke is determined by the winding angle of the contour curve of the revolving piston or cylinder and the positions of the suction port and the discharge port. When the winding angle of the contour curve of the revolving piston or cylinder is set to 360 °, the rotation angle θc of the rotating shaft in the compression stroke can be set to 360 °, and θc <360 by moving the seal point of the suction port or the discharge port.゜ can be. However, θc> 360 ° cannot be achieved. For example,
The rotation angle θc = 375 ° of the rotary shaft in the compression stroke of the compression element shown in FIG. 11 can be changed to θc = 360 ° by changing the position and size of the discharge port.

This can be realized by enlarging the discharge port so that the working chamber 15a and the working chamber 15b communicate with each other immediately after the suction end state in FIG. By performing such a change, it is possible to reduce irreversible mixing loss that occurs due to the difference in pressure rise between the two working chambers that occurred when θc = 375 °. Therefore, the winding angle of the contour curve is determined by the rotation angle θ of the rotating shaft during the compression stroke.
It can be said that this is a necessary condition for determining c, but not a sufficient condition.

The embodiment described above, that is, FIG.
In the embodiment shown in (1), the closed type compressor in which the pressure in the closed casing 3 is maintained at a low pressure (suction pressure) has been described. However, the use of the low pressure type has the following advantages.

(1) Since the electric element 2 is hardly heated by the compressed high-temperature working gas and is cooled by the suction gas, the temperatures of the stator 2a and the rotor 2b are reduced, and the motor efficiency is improved and the performance is improved. Improvement can be achieved.

(2) With a working fluid compatible with the lubricating oil 12 such as chlorofluorocarbon, the pressure is low, so that the proportion of the working gas dissolved in the lubricating oil 12 decreases, and oil bubbling occurs in bearings and the like. And reliability can be improved.

(3) The pressure resistance of the sealed container 3 can be reduced, and the thickness and weight can be reduced.

Next, a type in which the pressure in the sealed container 3 is maintained at a high pressure (discharge pressure) will be described. FIG.
FIG. 6 is an enlarged cross-sectional view of a main part of a high-pressure hermetic compressor using a swirling fluid machine according to another embodiment of the present invention as a compressor. In FIG. 16, components denoted by the same reference numerals as those in FIGS. 1 to 3 are the same components and perform the same operations. In the drawing, reference numeral 7b denotes a suction chamber formed integrally with the main bearing 7 by a suction cover 10, which is separated from the pressure (discharge pressure) in the sealed container 3 by a seal member 16 or the like. Reference numeral 17 denotes a discharge passage communicating between the inside of the discharge chamber 8b and the closed container 3. The operation principle and the like of the displacement type compression element 1 are the same as those of the low pressure (suction pressure) type described above.

The flow of the working gas, as indicated by the arrow in the figure, is such that the working gas that has entered the suction chamber 7b through the suction pipe 13 passes through the suction port 7a formed in the main bearing 7 and has a positive displacement. 1, where the rotation of the drive shaft 6 causes the revolving piston 5 to perform a revolving motion to reduce the volume of the working chamber 15, thereby being compressed. The compressed working gas is
The discharge valve 9 is pushed up through the discharge port 8a formed in the end plate of the sub-bearing 8 to enter the discharge chamber 8b, to enter the closed container 3 through the discharge passage 17, and to be connected to the discharge container connected to the closed container 3. It flows out from a pipe (not shown).

The advantage of such a high pressure type is that lubricating oil 1
Since the pressure of the rotary shaft 2 is high, the lubricating oil 12 supplied to each bearing sliding portion by the centrifugal pump action or the like due to the rotation of the drive shaft 6 is supplied into the cylinder 4 through a gap or the like at the end face of the revolving piston 5. Therefore, the sealability of the working chamber 15 and the lubricity of the sliding portion can be improved.

As described above, in the compressor using the positive displacement fluid machine of the present invention, it is possible to select either the low pressure type or the high pressure type according to the specifications and use of the equipment or the production equipment, etc., and the degree of freedom in design is greatly increased. To expand.

Next, a revolving piston according to an embodiment of the present invention will be described. FIG. 17 is an explanatory diagram of this, and FIG.
It is sectional drawing of the revolving piston 5 in CC cross section in (a). FIG. 18 is a diagram showing the portion a in FIG.
It is the figure shown from A section direction. FIG. 17A is a portion that receives the rotation moment M in the revolving piston 5,
It is a part where the revolving piston 5 and the cylinder 4 slide.
Since this is near the suction port 7a, a working gas having a low temperature and a high oil viscosity is supplied, and an oil film in the sliding portion is easily secured.

However, if the revolving piston 5 is formed as an integral part, the sliding form is one-way sliding with intermittent contact and load applied, so that the wear is relatively large. The swivel piston according to the embodiment of the present invention includes:
In order to further secure the sliding reliability, a needle bearing is provided as a rolling element 5c at a position receiving the rotation moment M of the revolving piston, and a part of the outer peripheral surface of the revolving piston is formed by the outer peripheral surface of the outer ring of the needle bearing.

For this purpose, the revolving piston 5 is composed of a rolling element 5c, a core 5d and a holding plate 5e for attaching the rolling element 5c to the core 5d. Since the revolving piston in FIG. 17 has three threads, the rolling elements 5c are attached to three portions that receive the rotation moment M. With this configuration, the sliding form between the revolving piston 5 and the cylinder 4 is rolling, so that wear can be reduced as compared with sliding sliding, and sliding reliability can be further ensured.

FIG. 19 is a view for explaining another example of a method of mounting a needle bearing as the rolling element 5c.
FIG. 2 is a diagram showing a portion a in FIG. In the embodiment described with reference to FIG.
Although it is composed of three points c, a core material 5d, and a holding plate 5e, in the present embodiment, the revolving piston 5 is a two-part member to support the rolling element 5c. With this configuration, the number of parts can be reduced,
Cost can be reduced. Furthermore, by assembling the revolving piston 5 into a shape as shown in FIG. 20 (a) or (b), it becomes possible to assemble it more easily. FIG. 21 shows the cylinder 4 and the revolving piston 5,
FIG. 18 is a view for explaining still another embodiment, and is a cross-sectional view showing a cross section similar to FIG. 17 for the revolving piston 5 and the cylinder 4. In the configuration of the revolving piston in the embodiment shown in FIG. 21, not only three portions of the revolving piston 5 but also a part of the vane 4 b of the cylinder 4 receive the rotating moment of the revolving piston during the operation of the compressor. Therefore, three points of the revolving piston and three points of the cylinder vane
A needle bearing is mounted as a rolling element at this position in order to slide at six positions in total. in this case,
In the vane 4b of the cylinder 4, it is necessary to support the needle bearing by using a holding plate similar to the revolving piston described in FIG.

Next, the mounting position of the rolling element in the revolving piston according to the embodiment of the present invention will be described. FIG.
(A) is a sectional view of the revolving piston 5 in a CC section in FIG. 1 (a), and (b) is a view showing a part a in (a) from the AA section direction. In FIG. 22, FIG.
The needle bearing which is a rolling element is supported by the method shown in FIG. In FIG. 22A, the outer peripheral surface of the revolving piston 5 protrudes outside the outer peripheral surface of the outer ring of the needle bearing, and the interval t between the outer peripheral surface of the revolving piston 5 and the outer peripheral surface of the outer ring of the needle bearing is AA direction. Is the smallest.

In this embodiment, since the outer peripheral surface of the outer ring of the needle bearing is retracted from the outer peripheral surface of the revolving piston 5, the needle bearing does not contact the inner wall surface of the cylinder 4 at the beginning of the compressor operation. Therefore, when the conformable film 5f is formed on the outer peripheral surface of the revolving piston 5 by a method such as a phosphate conversion treatment and the film thickness is _Tst , the outer peripheral surface of the outer ring of the needle bearing and the revolving piston 5 The needle bearing is mounted at a position where the minimum distance t from the outer peripheral surface is 0>t> -T _st (the case where the outer peripheral surface of the revolving piston 5 protrudes from the outer peripheral surface of the rolling element 5c is minus (-)). . As a result, at the beginning of the operation of the compressor, the conformable film on the outer peripheral surface of the projecting revolving piston 5 slides and slides with the cylinder 4, so that the conformable film wears in a short time. When the outer peripheral surface of the rotating piston 5c comes into contact with the cylinder 4 and rolls and slides, it stops, so that further wear does not progress, and the outer peripheral surface of the revolving piston 5 at this time always has a conformable film 5f. Remains, so that there is no gap between the cylinder supporting the rolling element 5c and the cylinder, and good sealing properties are obtained.

Therefore, loss due to leakage does not occur. Furthermore, the conformable film 5f on the outer peripheral surface of the revolving piston 5 can absorb an assembly error or the like, thereby contributing to an improvement in compression performance and a reduction in noise. In this state, rolling contact with the needle bearing and sliding sliding due to the outer peripheral surface of the orbiting piston at the portion supporting the needle bearing coexist at the contact portion receiving the rotation moment M. Since the sliding is performed by the conductive film 5f, there is no risk of burning or the like, and the sliding reliability is not impaired.

Further, by forming a conformable film 5f on the outer peripheral surface of the revolving piston 5, the conformable film is worn at a portion where a load is applied at a contact portion between the revolving piston and the cylinder, so that the resilient film is worn at an unnecessary portion. , And the rotation moment M is ultimately received only by the needle bearing. Therefore, it is not necessary to provide the rolling elements in the cylinder 4, and even if the revolving piston has a shape as shown in FIG. 20, needle bearings may be provided at three places on the revolving piston side. Therefore, simplification of assembly, reduction of the number of parts, and cost reduction can be achieved.

In FIG. 22, a needle bearing is provided as a rolling element on the side of the revolving piston 5, but the same effect can be obtained by providing a needle bearing on the side of the cylinder 4 which receives the rotation moment M. Although the needle bearing is supported by the method shown in FIG. 18, the same effect can be obtained by the method of forming the revolving piston by using two members as shown in FIG. 19 and the method shown in FIG. Demonstrate.

In the above-described configuration of the revolving piston and the cylinder, a needle bearing is used as a rolling element in the sliding portion. However, the effect is the same even if the revolving piston is configured with a bearing of another shape depending on the size of the revolving piston. is there.

In the above configuration, the rolling elements do not have to be bearings. FIG. 23 is a diagram illustrating an example in which a rolling member is used as a rolling element. FIG. 23 shows a portion of the orbiting piston corresponding to the portion a in FIG. In (a), a cylindrical rolling member is used. In this case, since the contact time of the outer wall surface of the rolling element 5c with the inner wall surface of the cylinder is short, it is necessary to mount the rolling member 5c at a position where the rotation moment M comes into contact with the maximum, but without using a rolling bearing or the like. Therefore, the assembling is easy and the cost can be reduced.

(B) is a structure for solving the problem of contact time, which is a problem in (a). The contact time is the same as that when a rolling bearing is used by using a rolling element having the illustrated shape. It can be. In this case, the method of supporting the rolling element 5c is shown in FIGS.
This is the method (a). In this case, the rolling element does not need to be an integral part, and the constituent parts are separated so that the structure shown in FIG.
It is also possible to adopt a method as shown in FIG.

FIG. 24 shows an air conditioning system to which the positive displacement compressor of the present invention is applied. This cycle is a heat pump cycle capable of cooling and heating, and includes the positive displacement compressor 30, the outdoor heat exchanger 31, and the fan 3 of the present invention described with reference to FIG.
1a, expansion valve 32, indoor heat exchanger 33 and fan 33
a, a four-way valve 34. An alternate long and short dash line 35 indicates an outdoor unit and 36 indicates an indoor unit.

The positive displacement compressor 30 operates in accordance with the principle of operation shown in FIG. 2, and starts up the compressor to cause a working fluid (for example, Freon HCFC22, R407C, R410A, etc.) between the cylinder 4 and the revolving piston 5. Is performed.

In the cooling operation, the compressed high-temperature and high-pressure working gas is supplied from the discharge pipe 14 to
Flows into the outdoor heat exchanger 31 through the
1a, the heat is radiated and liquefied by the blowing action, and is throttled by the expansion valve 32.
After being adiabatically expanded to become low temperature and low pressure, the indoor heat exchanger 33 absorbs indoor heat and is gasified, and then is sucked into the positive displacement compressor 30 through the suction pipe 13. On the other hand, in the case of the heating operation, as shown by the solid line arrow, the air flows in the opposite direction to the cooling operation, and the compressed high-temperature and high-pressure working gas flows from the discharge pipe 14 through the discharge pipe 14.
Flows into the indoor heat exchanger 33 through the
The heat is radiated into the room by the blowing action of 3a, liquefied, and the expansion valve 32
After being adiabatically expanded and adiabatically expanded to a low temperature and low pressure, the heat is absorbed from the outside air by the outdoor heat exchanger 33 to be gasified, and then sucked into the positive displacement compressor 30 through the suction pipe 13.

FIG. 25 shows a refrigeration system equipped with the positive displacement compressor of the present invention. This cycle is a cycle dedicated to freezing (cooling). In the figure, 37 is a condenser, 37
a is a condenser fan, 38 is an expansion valve, 39 is an evaporator, 39
a is an evaporator fan.

When the positive displacement compressor 30 is started, the working fluid is compressed between the cylinder 4 and the revolving piston 5, and the compressed high-temperature and high-pressure working gas flows from the discharge pipe 14 as shown by the solid arrow. After flowing into the condenser 37, the heat is radiated and liquefied by the blowing action of the fan 37a, throttled by the expansion valve 38, adiabatically expanded to a low temperature and low pressure, and endothermic gasified by the evaporator 39. After that, it is sucked into the positive displacement compressor 30. Here, since the displacement compressor of the present invention is mounted in both of FIGS. 24 and 25, a highly reliable refrigeration / air-conditioning system having excellent energy efficiency, low vibration and low noise can be obtained. Although the low-pressure type compressor 30 has been described as an example here, the high-pressure type compressor also functions similarly and can achieve the same effect.

[0098]

As described above in detail, according to the present invention, two or more working chambers are provided around a drive shaft, and rotation of each working chamber from the end of suction to the end of discharge is performed. For a compressor configured so that the rotation angle of the shaft is approximately 360 °, the rotation moment acting on the displacer is reduced by the rolling elements, and the friction loss and the leakage loss between the displacer and the cylinder are reduced. Accordingly, a positive displacement type fluid machine with improved performance and high reliability can be obtained. In addition, by mounting such a positive displacement fluid machine in a refrigeration cycle, a highly reliable refrigeration / air-conditioning system with excellent energy efficiency can be obtained.

[Brief description of the drawings]

FIG. 1 is a vertical sectional view and a plan view of a compression element of a hermetic compressor in which a positive displacement fluid machine according to the present invention is applied to a compressor.

FIG. 2 is a diagram illustrating the operation principle of the positive displacement fluid machine according to the present invention.

FIG. 3 is a longitudinal sectional view of a positive displacement fluid machine according to the present invention.

FIG. 4 is a diagram showing a contour configuration method of a displacer of the positive displacement fluid machine according to the present invention.

FIG. 5 is a diagram showing a contour configuration method of a cylinder of the positive displacement fluid machine according to the present invention.

FIG. 6 is a diagram in which the displacer and the cylinder shown in FIGS. 4 and 5 are superimposed.

FIG. 7 is a characteristic diagram showing a change in volume of a working chamber in the present invention.

FIG. 8 is a diagram showing a change in gas compression torque in the present invention.

FIG. 9 is a diagram showing a relationship between a rotation angle of a rotation shaft and a working chamber in a four-row wrap.

FIG. 10 is a diagram showing a relationship between a rotation angle of a rotation shaft and a working chamber in the three-row wrap.

FIG. 11 is a view for explaining the operation when the winding angle of the compression element is larger than 360 °.

FIG. 12 is a view for explaining an enlargement of a winding angle of a compression element.

FIG. 13 is a view showing a modification of the positive displacement fluid machine shown in FIG. 1;

FIG. 14 is a diagram illustrating a load and a moment acting on the displacer of the present invention.

FIG. 15 is a characteristic diagram showing a relationship between a rotation angle of a rotation shaft of a compression element and a rotation moment ratio.

FIG. 16 is a vertical sectional view of a main part of a hermetic compressor according to another embodiment of the present invention.

FIG. 17 is a sectional view of a displacer according to the present invention.

18 is a partial vertical sectional view of the displacer shown in FIG.

FIG. 19 is a partial longitudinal sectional view of another embodiment of the displacer shown in FIG. 17;

20 is a partial perspective view of yet another embodiment of the displacer shown in FIG.

FIG. 21 is a sectional view of a displacer and a cylinder according to the present invention.

FIG. 22 is a diagram of another embodiment of the displacer according to the present invention.

FIG. 23 is a partial perspective view showing another embodiment of the displacer according to the present invention.

FIG. 24 is a diagram showing an air conditioning system to which the positive displacement compressor of the present invention is applied.

FIG. 25 is a diagram showing a refrigeration system to which the positive displacement compressor of the present invention is applied.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 ... Volumetric compression element, 2 ... Electric element, 3 ... Airtight container, 4
... Cylinder, 4a ... Inner peripheral wall, 4b ... Vane, 5 ... Displacer (rotating piston), 5a, 41c ... Bearing, 5b
... Through holes, 5c ... rolling elements, 5d ... core materials, 5e ... holding plates, 5f ... conformable films, 6 ... drive shafts, 6a ... crank parts, 7 ... main bearings, 7a, 42a, 42b ... suction ports,
Reference numeral 8: auxiliary bearing, 8a: discharge port, 8b: discharge chamber, 9: discharge valve, 10: suction cover, 11: discharge cover, 12: lubricating oil, 13: suction pipe, 14: discharge pipe, 15, 4
5 ... working chamber, 16 ... seal member, 17 ... discharge passage, 18
... Processing jig, 18a ... Base, 18b ... Pin part, 18c
... Clamp, 19 ... Working tool, 19a ... Grinding tool, 19
b: cutting tool, 30: positive displacement compressor, 31: outdoor heat exchanger, 32, 38: expansion valve, 33: indoor heat exchanger, 34 ...
4-way valve, 37: condenser, 39: evaporator, 40: fixed member, 40a: fixed spiral member, 40b: end plate portion, 40c: main bearing portion, 41: revolving member, 41a: revolving spiral member, 41b
... Reinforcement plate, 42 ... Ring part, 42a ... Suction chamber, 43 ... Check valve, 44 ... Shaft sealing device, o ... Displacer center, o '
... Center of cylinder, Om ... Center of turning side member, Of ... Center of fixed side member.

 ──────────────────────────────────────────────────続 き Continuing on the front page (72) Inventor Hiroaki Hata 800 Tomita, Ohira-cho, Ohira-cho, Shimotsuga-gun, Tochigi Pref.

Claims (1)

[Claims] 1. A cylinder having an inner wall formed between a pair of end plates and having a continuous curved shape in a plane, and an outer wall provided so as to face the inner wall of the cylinder. A displacer forming a plurality of spaces by an outer wall and the end plate, and the space present along the contour of the displacer is alternately a space in a suction stroke and a space in a compression or discharge stroke. In the displacement type fluid machine in which the cylinder inner wall and the displacer outer wall curve are formed, a part of an outer wall surface of the displacer which is a contact point when a space is formed by the cylinder inner wall and the displacer outer wall is formed on an outer peripheral surface of a rolling element. A positive displacement fluid machine that is formed and the contact at the contact point is a rolling contact.